Fluid pressure bearing



P. L. GERARD FLUID PRESSURE BEARING April 7, v1953 Filed-Jan. 26, 1951 2 n 1b 2b :n 4o sb o 1n ab 9o ma n'a no 13o :4o 15o fsa/10. fsu

SHEETS-SHEET 1 In venan'- Patented Apr. 7, 1953 PRESSURE BEARING Paul LGrard, Paris, France Application InuaYfZG, 1951, ySeliztlrND.2022; 869v In France `January 31, 1950 of said fluid is so determined as to provide a safety -factor of about 5, for example, so that all possible dynamic effects are 'taken into .'acceunt.

`Assuming, e. g.,1a rotor Weight of l20 kg. acting upon each bearing, .such a bearing should be so designed as to provide areaction of 100 kg. With this factor of safety, there will be 'no risk of accidental failure in the case bf a Anormally .balanced rotor.

VIn certain applications of a bearing fof the type f described, however, there is a risk due Vto `unexpected causes, of the rotor being subjected .to a rotating force resulting Ifrom fa Weight unbalan'ce thereof. v'1h`us,.for -example, in the :case 4of a shaft journalled in bearings of the type described and carrying a turbine rotor, a .blade of the rotor may bev broken, which results in immediately throwing the rotor out of balance, so that the same issubjected to la considerable rotating force, the value fof which is Y m being 'the mass of the detached blade, `n the angular velocity Yof the rotor, and r the distance from the rotation axis to the centre 'ofggravity of the detachedmass.

Since the value 'of 'the rotating ior'ce is a 'fu-nction of the 'square of angular velocity, vvsaid 'force is rapidly increasing With high 'rotation speeds of the shaft,jwnicn are precisely 'to be "en-countered in turbines. Thus, the above `mentioned rotating 'force lmay happen to exceed materially `the reaction Vforce which can'be normally 'provided by the bearing. There is 'thus a risk of the shaft coming into friction 'contact Wit-h the bearing,

' which would result in a rapid Wear of `said shaft and bearing and even in causing serious vaccidents.

' An oblie'ct of my .invention `is to provide 'a bearing oi thetyp'e described, inw'hicn all above men 'In bearings of the Vtype described, the pressure 5 2 tioned drawbacks are overcome and, more ,particularly, in which the 'effects of any 'unexpected offsetting of the rotor are .automatically suppressed.

Due to the very fact that 'the .rotating force. when the rotor is out of balance, 'is .a function of the square of the angular velocity o'fsaid rotor, it is indispensable, 'if the reaction s-'to be sufricient for any rotation speed, to 'take into account the nature of the rotating force.

Another object of my invention is to provide for this purpose a bearingof the type described, in which non-return valves .are .interposed in the feeding path of the bearing chambers, said nonreturn Valves being so designed that in normal conditions the feeding takes place ythrough said valves, which remain open under the `action cf the feeding pressure, Whileasaid valves vare automatically closed, in predetermined `conditions under the action of the rotating force of a shaft which has been unexpectedly offset.

I have foundthat, with @the device according to my invention, the Lreaction .forces developed are not only considerable, but .also ,proportional to thelsquare of the angular velocity of therotor, so that, Whatever may .be the value of -said angular velocity, the reaction forces are suicient to balance said rotor.

Fig. 1 represents, in diagrammatic form, a cross-sectional View of 'a bearing having a shaft supported therein for rotation, the clearance being greatly exaggerated for purposes of illustra tion;

Fig. 2 represents a curve of pressure variations in function of the angle of inclination of the rotating force to the bearing axis;

Fig. 3 represents, diagrammatically, a shaft carrying a turbine rotor and iournalled `in two bearings;

Fig. 4 represents, in somewhat diagrammatic form, a cross-sectional view o'f a bearing according to the invention, the clearance and certain other dimensions being exaggerated for :purposes of illustration, and

Fig. '5 'represents a fdetail @axial section :through a modified form of Valve.

Assume 'in a `iiuid 'bear-ing nf centre V'O ,(Fig. 1) a shaft fof centre O surged lby :a rotatingiorce OF running through O. Let `OX be fa, .reference diameter, stationary with .respect Eto the bearing. Let U=wt, be the angle .FOX at .a igiven instant, w being the angular velocity `of the shaft. .Let nally a be the distanceb'etv/.een and @1, si. fe. the offsetting of the `shaft fdue to force at fsaid =instant. 1

Considering a chamber c, having OX for its axis and the angular size of which is reasonable, it may be said that clearance :c in front of said chamber at a given instant, is:

m=b+a cos wt b being a constant which is equal to half the diametral clearance between the shaft and the bearing. Volume V of the fluid contained in said chamber is, neglecting a constant:

S being the area of the chamber.

This may be re-written:

V=S` b+a cos wt) The variations of said volume V are, neglecting the sign:

dV a awS sin wt From a small value of U upwards to a value nearly amounting to 180 computations effected in well defined specific cases have shown as described hereunder that the pressure in the chamber is higher than the feeding pressure. In other words, the non-return valve through which the feeding of the chamber takes place is closed. The volume variations thus correspond to a leakage flow occurring solely through clearance between the shaft and bearing along perimeter P of the chamber.

I have found that said leakage takes place through eddy ow, i. e. its rate is proportional to the square root f p, p being the pressure in the relevant chamber, Thus the rate of leakage at a given instant is equal to KPVpm, or KPVpUi-t-a cos et) K being a coeflicient which takes into account the contraction effect as the fluid passes through the clearance, the specific mass of the fluid and the kind of units used. Since this rate of leakage is equal to the volume variation versus time, i. e.

lli

the following equation may be written:

awS sin wt=KP\/p(b+a cos wt) and thence or, substituting U to et:

@ awS sin U KPB 1 -i-- cos U) Coeicients P and S may be easily computed for any specific bearing.

Coefcient K is determined by choosing a well defined fiuid.

Assuming a predetermined clearance between the shaft and bearing, the value of coefficient b may also bedetermined; finally, it will be assumed that, the shaft, under the action of an indefinite rotating force, rotates with a constant offsetting a. It is thus clear that, for each value of angle U, the pressure is proportional to the square of .angular velocity w. Y

Now, as mentioned above, the rotating force .4 resulting from a weight unbalance of the shaft is also proportional to the square of angular velocity. Therefore, whatever may be the rotation speed, the reaction offered by the bearing remains proportional to the rotating force, without any variation of the shaft offsetting under the action of said rotating force.

A specific example is given hereunder as an illustration.

Let the diametral clearance between the shaft and bearing be 40 p,

The shaft offsetting under the action of the rotating force 10 p,

The rotation speed of the shaft 40,000 R. P. M.,

The inner diameter of the chamber 30 mm. and the pressure of the feeding uid (water) 1 kg./cm.2.

In Fig. 2 is plotted a curve of the variations of pressure p in the chamber, versus angle U by which the rotating force is inclined on axis OX of the bearing. It is to be noted that:

p=1 kg./cm.2 for U=about 19 (closing of the valve) and for U=about 174 (re-opening of the valve).

On the other hand, p is a maximum for U= and reaches then a value of about 28 kg./cm.2.

Assuming that in normal running conditions, the bearing is capable of generating a pressure equal to half the feeding pressure, which is 1 kg./cm.2, it will be seen that the maximum overpressure to be generated by the bearing, when vibrating, reaches, thanks to my invention, about 50 times the normal pressure.

Referring to these drawings, there are shown at l and 2, two bearings in which is journalled a shaft 3 carrying the rotor 4 of a turbine.

Bearing I is shown in detail in Fig. 4. In the example shown, this bearing comprises four feeding chambers 5, fed with a pressure fluid through throttlings 6, and separated by grooves l, adapted to discharge the fluid escaping from said chambers 5 into a low pressure space.

The centre of the bearing is shown at O. It will be assumed that shaft 3 journalled in this bearing has O' for centre and is subjected to a rotating force F as explained above. Under the action of said force, the shaft tends to come into contact with the bearing at a point M aligned with F, force F extending as shown along a diametral direction OX.

It is clear that point M rotates with F as well as with the shaft.

Observing the fluid Volume in any chamber, e. g. chamber 5c of Fig. 4, it will be found that said volume is subjected to substantially sinusoidal periodical variations. Said volume is a minimum when point M corresponding to the minimum distance between the shaft and bearing is in the position shown in the drawing. Said volume is a maximum when point M is diametrically opposed to the above mentioned position. The pressure of the fluid in the chamber increases as the volume decreases and conversely. If the value of said pressure exceeds that of the feeding pressure, the uid is forced back through throttling Bc and the maximum value of the pressure in chamber 5c is then determined by the leakage along the perimeter of the chamber into the low pressure space added with the leakage through the throttling into that space which is subjected to the feeding pressure.

If, however, there is interposed, according to the invention, in the feeding path of each chamber, a non-return valve constituted, in the ex-.

`shown located upstream the throttlings.

ample shown in Fig. 4, by a ball t urged towards its seat 9 by a spring l0, the maximum value of the pressure in the chamber is solely determined by the leakage along the perimeter of said chamber, any back flow towards the feeding duct being opposed by the non-return valve. As a result, the value of the pressure in chamber 5c will be materially higher than the pressure which would be experienced ifthe non-return valve according to the invention were omitted.

The device according tothe invention operates as follows: n

1. In normal operating conditions, all nonreturn valves are open and the feeding of the chambers takes place through the throttlings. The bearing, which operates as described in detail in the above cited patent, is capable of balancing the normal forces which are transmitted thereto from the shaft as long as said forces do not exceed a certain value.

2. In accidental operating conditions, the throttling corresponding to a given chamber feeds the same during substantially the half revolution when the Shaft tends to be separated from said chamber; on the contrary, during the following half revolution, when the shaft tends to be brought closer to the chamber, the non return `valve is re-closed and the shaft is subjected to a considerable resistance.

Thus, the reaction of which the bearing is capable is several times higher than in normal running condition.

It will be easily understood that the system described above may be used with all types of chamber-and-groove bearings described in the i cited patent, whether male or female and also with bearings of the type in which each chamber is fed not directly from the pressure source but through a diametrically opposed groove. In each case, to provide a bearing according to the invention, it sui'lices to interpose a non-return valve in the fluid intake of each feeding chamber.

In Fig. 4 the anti-return valves have been It would be possible as Well'to dispose said valves downstream the same. The arrangement shown in Fig. 4 is, however, to be preferred since, in

normal running conditions, the fluid flowing through a valve undergoes a pressure fall Ap corresponding to the strength of the spring.

On the other hand, the fluid undergoes another pressure fall when ilowing through the throttling. The relative value of pressure fall Ap is therefore lower when taking place upstream the throttling.

A particular object of the invention is therefore to provide a device of the type described in which the non-return valve of each feeding chamber is disposed upstream the corresponding throttling.

The spring strength has a critical value. With a spring stronger than this critical value, the

6 pressure fall Ap would be too considerable, which would result in an excess of feeding pressure and thence of the pumping power. Conversely, with a spring weaker than said critical value, the valve would be closed too late, particularly with an unbalanced rotor revolving at a high speed. The spring should be therefore carefully calibrated. The valve proper is preferably as light as possible.

As a specific example, at a rotation speed of about 30,000 R. P. M. with a shaft of about 30 mm. diameter, the following values have been found satisfactory: sheet-steel valve 0f 5.5 mm. diam-V eter and a 0.2 mm. thick, spring strength of about 50 g. e A valve designed according to these conditions has been shown in Fig. 5. There is shown at Il the valve body in the bore of which is housed a small platei2 urged towards a valve-seat 9 by a spring l0. The throttling is provided in the end of body Il opposed to seat 9. Seat 9 may be secured in the bore of body Il, by soldering or crimping, e. g. by turning over a projecting edge i3.

What is claimed is:

1. In a bearing member having a bearing surface for supporting a movable element with an annular clearance space therebetween, a plurality of circumferentially spaced pressure chambers, means for separately conducting fluid under pressure into each of said pressure chambers including separate ducts, longitudinal evacuation grooves between said pressure chambers through which said fluid is discharged therefrom, and a non-return valve in each of said ducts, in close proximity to each said pressure chamber, each valve being adapted to be closed automatically whenever the fluid pressure in the corresponding pressure chamber exceeds that of the fluid conducted thereto.

2. A bearing device according to claim 1, in which each said duct includes a constricted passage and in which each said non-return valve is interposed in each said duct upstream said constricted passage.

3. A bearing device according to claim 1, in which each said valve includes an obturating member in the form of a thin sheet-metal disc.

4. A bearing device according to claim 2, in which each said valve includes an obturating member in the form of a thin sheet-metal disc;

PAUL L. GRARD.

REFERENCES CITED The following references are of record in the file of this patent:

UNITED STATES PATENTS Number Name Date 1,906,715 Penick May 2, 1933 2,254,670 Turner Sept. 2, 1941 2,495,516 Foster Jan. 24, 1950 

